Automatic distortion control for gear type pumps and motors

ABSTRACT

An internal gear pump or motor in which distortion of the ring gear is controlled so that, under operating pressure, the ring gear becomes elliptical in configuration with the major diameter of the ellipse being at right angles to the neutral axis of the device. Mechanical, hydraulic, and combined mechanical-hydraulic means are disclosed for controlling distortion of the ring gear.

United States Patent lnventor Robert Wesley Brundage St. Louis. Mo.

Appl. No. 851,633

Filed Aug. 20, 1969 Patented June 8, 1971 Assignee Emerson Electric Co.

St.L0uis, Mo.

AUTOMATIC DISTORTION CONTROL FOR GEAR TYPE PUMPS AND MOTORS 15 Claims, 10 Drawing Figs.

U.S. Cl 418/108, 418/171 Int. Cl F01c 19/00, F04c 27/00, F010 1/10 Field of Search 418/71,

104, 107, 108, 171; 103/126 A, 126 BA; 230/141 References Cited UNITED STATES PATENTS Nichols Bunte Witchger Brundage Brundage Ross Primary Examiner-Carlton R. Croyle Assistant Examiner-John J. Vrablik AttorneyMeyer, Tilberry and Body 103/126A 103/126A 103/126A 103/1 20C 103/126A 103/126A ABSTRACT: An internal gear pump or motor in which distortion of the ring gear is controlled so that, under operating pressure, the ring gear becomes elliptical in configuration with the major diameter of the ellipse being at right angles to the neutral axis of the device. Mechanical, hydraulic, and combined mechanical-hydraulic means are disclosed for controlling distortion of the ring gear.

PATENTEU JUN 8 |97l SHEET 1 OF 2 E m mD N N EU R mB W m E B O R PATENTEDJUN 8l97l 3583839 SHEET 2 OF 2 FIG.9

INVEN R. ROBERTWBRU AGE 60 fyw 650% ATTORNEYS AUTOMATIC DISTORTION CONTROL FOR GEAR TYPE PUMPS AND MOTORS This invention relates to the field of hydraulic pumps or motors and, more particularly, to means for controlling distortion of the ring gear in internal gear pumps or motors of the type commonly known as gerotors.

Internal gear-type pumps and motors are well known in the art. Specific examples of such pumps or motors may be found in my prior U.S. Pat. Nos. 2,956,512, 3,007,418, 3,034,446, 3,034,447, 3,034,448, 3,127,843, 3,188,969, 3,198,127, 3,240,158 and 3,427,983. In general, these internal gear-type pumps or motors consist of a pair of gear-shaped elements, one within the other, mounted in a fluid chamber. The externally toothed gear, or pinion rotor, is supported on a shaft rotatably mounted in the housing. The internally toothed gear, or ring gear, is mounted for rotation on an axis spaced from that of the shaft. The tooth form of each gear mates with the other in such a way that each tooth of the pinion gear is always in sliding sealing contact with the inner surface of the ring gear. Sealing members engage the axial faces of the gear members so that, as the gear members rotate, they define a plurality of closed chambers revolving about the axis and progressively increasing to a point of maximum volume which corresponds to the point of open mesh of the gears and then to a point of minimum volume which corresponds to the point of closed mesh of the gears. Normally, the chambers which are decreasing in volume communicate with the discharge port and are at high hydraulic pressures while the chambers which are increasing in volume communicate with an inlet port and are at relatively low pressures. A line drawn through the points of minimum volume and maximum volume define what may be termed a neutral plane or axis of eccentricity. As is well known, at the closed mesh position the relative velocity between the two gears is very low and the drive angle is such that the pinion gear drives the ring gear. It is also well known that the gear teeth, particularly at the open mesh position, should be in sealing engagement with each other thereby to separate the fluid pressures in the increasing volume chambers on one side of the neutral axis from the fluid pressures in the decreasing volume chambers on the other side of the neutral axis.

THE PRIOR ART PROBLEM One problem that has been experienced with the prior-art devices is the effect of the hydraulic pressures on the internally toothed ring gear. This gear, which normally is of relatively lightweight construction, is exposed to relatively high pressures, e.g., 2,000 pounds per square inch over a substantial portion or are (approximately 16x0") of its inner surface while an opposite and equal portion or are of the inner surface is exposed to relatively low pressures. These large pressure differentials frequently result in distortion of the ring gear to an elliptical or egg-shaped configuration. The reason for this is apparent from a force analysis of a typical operating cycle. Thus, the high hydraulic pressures on the high-pressure are force the ring gear against its bearing surface which, of necessity, must be slightly larger than the ring gear so as to provide a running clearance Accordingly, the surfaces engage at the point of tangency. These high hydraulic pressures may be considered as resisted by a single force which acts against the exterior of the ring gear on the point of tangency which is on a line approximately perpendicular to and on the high-pressure side of the neutral axis. The high pressures, however, exert uniform forces over a substantial are on both sides of this line which forces tend to enlarge the are on this side of the gear forcing the gear into an elliptical egg-shaped configuration with the major diameter of the ellipse lying along the neutral axis of the hydraulic device.

This tendency of the ring gear to become elliptical when under pressure has two specific adverse effects on the operation of the hydraulic device. Thus, the teeth of the two gears, both at the open mesh and closed mesh positions, tend to become separated as the ring gear distorts radially outward. At the open mesh position, the result of the tooth separation is that the sealing engagement of the teeth is destroyed with consequent leakage between the adjacent pressure chambers. At the closed mesh position, the separation of the teeth results in a poor drive angle between the teeth of the inner and outer gears so that the desired driving force between the two gears does not occur.

SUMMARY OF THE INVENTION This invention remedies the above-described problem by controlling the distortion of the ring gear either in terms of preventing distortion or distorting the ring gear in the opposite direction to improve the operation of the hydraulic device. In its preferred embodiment, the invention achieves this result by employing a combination of forces to force the ring gear to an elliptical shape in which the major diameter of the ellipse is substantially at right angles to the neutral axis of the device.

It will be appreciated that with the major diameter of the ellipse at right angles to the neutral axis of the gears, the teeth of the gears at the open mesh and closed mesh positions not only do not tend to separate but, rather, are brought into even closer contact thereby assuring both a good drive angle between the teeth at the closed mesh position and a good sealing relationship between the teeth at the open mesh position.

This result is achieved in a preferred embodiment of the invention by employing a combination of hydraulic and mechanical forces to distort the ring gear to the desired elliptical configuration. More specifically, the hydraulic device employs, in addition to the inner and outer gears, a bearing member having a cylindrical bore in which the ring gear is received. The bearing member, in turn, is supported within a bore in the housing of the device. The bearing member is subjected to external forces so as to cause the bearing member to distort to an elliptical configuration with the major diameter of the ellipse being at the desired right angles to the neutral axis of the device. As the bearing member becomes elliptical, it exerts a force on the right gear to distort it to a like elliptical configuration.

The external forces may be purely mechanical or may be supplemented by hydraulic pressures introduced at strategic places around the bearing surface between the ring gear and the bearing member.

Alternatively, the elliptical configuration 0f the ring gear may be controlled solely by hydraulic pressures.

Referring now to the drawings, wherein like reference numerals indicate like parts in the various views:

FIG. 1 is a cross-sectional view of an internal gear-type hydraulic device illustrating a preferred embodiment of the invention.

FIG. 2 is a sectional view taken along lines 2-2 of FIG. 1.

FIG. 3 is a sectional view taken along lines 3-3 of FIG. 1.

FIGS. 4 and 4a are schematic illustrations of the forces acting on the ring gear under normal operating conditions.

FIG. 5 is a plan view of the special bearing member.

FIGS. 6 and 7 are views similar to FIGS. 2 and 3 illustrating a modified form of the invention.

FIG. 8 is a schematic illustration of hydraulic forces acting on the ring gear in accordance with the principles of this invention.

FIG. 9 is a schematic illustration of the deformation of the ring gear under the forces of FIG. 8.

Referring now more in detail to the drawings wherein the showings are for the purpose of illustrating a preferred embodiment of the invention only and not for the purposes of limiting same, FIG. 1 illustrates a hydraulic device, indicated generally by the reference numeral 10, having an internal cavity or fluid chamber 12 defined by end walls 14 and 16. Disposed within the cavity 12 is a pair of gear members which, for the purpose of the description of this invention, comprise a hydraulic pump. These gear members comprise an externally toothed inner gear (or pinion rotor) 18 and an internally toothed gear member (ring gear or ring rotor) 20. The inner gear 18 is keyed to a shaft 22 rotatably supported within the housing for rotation about an axis 24. The internally toothed gear member is supported for rotation about an axis 26 spaced from the axis 24. The support for the gear member 20 comprises a bearing member 28 having a cylindrical bore 30 within which the gear member 20 is received.

As is conventional in devices of this type, the gear member 20 has one tooth more than that of the gear member 18. The teeth of the inner gear 18 are intended always to be in sliding sealing engagement with the inner surface of gear member 20, in a manner well known in the art. Thus, as the gears rotate, the teeth cooperate to form a plurality of fluid chambers 32H and 32L which progressively increase and decrease in volume. A line drawn through the point of minimum volume and the point of maximum volume define the neutral plane or axis 34 of the device. As is well known, the point of minimum volume comprises the closed mesh position of the gears while the point of maximum volume comprises the open mesh condition of the gears.

The walls 14, 16 are each in sliding sealing engagement with the axial faces of the gears 18 and 20 thereby to define a closed fluid chamber internally of the gears. The wall 16 has arcuate extending inlet and outlet manifold ports 40, 42, respectively, formed therein and spaced from the axis of rotation so as to communicate with the increasing and decreasing volume chambers. The inlet manifold port has a supply passage 43 through which a source of fluid may be introduced and the outlet manifold port 42 has a passage 44 through which fluid may be discharged. The spaces 45 on the wall 16 between the arcuate ends of the manifolds 40, 42 function as lands to prevent communication from one manifold port to the other.

THE DISTORTION PROBLEM A graphic analysis of the forces generated in the device described so far is helpful in understanding the distortion problem solved by this invention.

Considering the hydraulic device as a pump, and assuming a clockwise rotation of the gear 18 as viewed in FIG. 2, the chambers 32H on the left side of the neutral axis 34 are all at high hydraulic pressures. These pressures exert a uniform radially outward force over an arc of slightly less than one-half of the entire inner surface of the ring gear 20. All of the forces fmay be integrated and considered as a single large, radially outward load or resultant force P which acts approximately perpendicular to and on the high-pressure side of the neutral axis 34.

On the opposite side of the neutral axis 34 is the low pressure or inletpressure of the device. Between the low'pressure and high-pressure arcs is a pressure gradient with points of one-half pressure falling approximately along the neutral axis 34.

The journal bearing surface 30 on which the ring gear 20 is rotatably supported has a normal clearance relative to the outer surface of the ring gear. For example, the bearing clearance may fall within the range of 0.0005 to 0.00l inch per inch of diameter. with these normal bearing clearances, a certain small but finite shifting ofthe ring gear within the journal bearing is to be expected because of the unbalanced pressures on the inner surfaces of the ring gear.

In the normal operation of this device, and assuming normal bearing clearances, a force analysis would indicate that the forces on the ring gear are essentially as depicted in FIG. 4 wherein the bearing clearance is greatly exaggerated. Thus, the high pressure of the device, represented by the force P, exerts a radially outward force against the inner periphery of the ring gear 20. Since the pressures in the device are not in balance, the force P causes the gear 20 to shift within the bearing clearance in the direction of the unbalance until a counterbalancing force Pj is established by the lubricating film of the journal bearing. This force Pj is equal to the summation of forces f of the parallel components of the forces created by the hydraulic pressures. In this condition, the two forces P, Pj are counterbalanced and the ring gear reaches a point of equilibrium within the journal bearing. However, as pointed out above, the entire are on the high-pressure side of ring gear 20 is exposed to the high-pressure forces f. As is apparent from FIG. 4a, the forces f act on either side of the point against which Pj acts, thereby deforming the ring about the point Pj as a fulcrum point. The result is that the ring gear 20 becomes elliptical with the major axis of the ellipse falling along the neutral axis 34.

Moreover, assuming the point of low pressure is spaced approximately l from the point of high pressure P, it follows that there are opposed points of one-half pressure which fall along the neutral axis 34 of the device. These points of onehalf pressure, represented by the arrows l/2P in FIG. 4, act in opposite directions radially outwardly against the inner periphery of the ring gear 20. Since the ring gear 20 has shifted within the bearing clearance, it will be apparent that the ring gear is no longer uniformly supported on the bearing surface 30 but instead is supported on the line of tangency of the two surfaces. As a result the forces l/2P acting radially outward are not fully opposed by the journal bearing and will tend to distort the unsupported adjacent portions of ring gear 20 radially outwardly. Thus, the forces l/2P also force the ring gear to assume an elliptical configuration in which the major diameter of the ellipse falls along the neutral axis 34.

As this distortion of the ring gear occurs, the teeth on the ring gear 20 tend to separate from the teeth on the pinion gear 18 at the open mesh position and thus destroy the sliding sealing engagement of the teeth. The teeth at the closed mesh position also tend to separate and, as pointed out above, a poor drive angle between the teeth results from this separation.

In accordance with the principal aspects of this invention, means are provided for generating forces which control the distortion of the ring gear and which prevent any tendency of the gear teeth to separate either at the open mesh or the closed mesh positions.

M ECHANICAL DISTORTION CONTROL Various mechanical means may be employed to control distortion of the ring gear. One such arrangement is illustrated in FIGS. 2 and 5 wherein the bearing member 28 is employed for accomplishing this objective. The bearing member 28 is received in a cylindrical bore 50 within the housing with the basic or unstressed outer diameter 51 of the bearing member 28 being greater than the diameter of the bore 50. However, the outer surface of the bearing member 28 has relieved portions 52 ofa reduced diameter which extend through an arc of at least 200 thereby permitting reception of the bearing member 28 within the bore 50 when the outer diameter 51 is reduced by compressing the bearing member 28 to an ovate shape with the major diameter generally at right angles to the neutral axis. The relationship between the bearing member the bore 50 is such that the bearing member 28 engages the inner surface of the bore 50 at two points 54, 56. The portion of the outer periphery of the bearing member between these two points is spaced from the bore 50.

The positioning ofthe points 54, 56 within the bore 50 is important. Thus, as is apparent from FIG. 2, both points 54, 56 lie on the high-pressure side of the neutral axis 34 with the resultant force P acting approximately midway between the two points. In this manner, the forces generated on the bearing member 28 by the resultant force P are applied to a specific section of the bearing member. By controlling the section modulus of the bearing member 28, it will be apparent that distortion of the bearing member can be controlled, both in terms of amount of distortion and the configuration of the distortion. Since the points 54,56 engage the periphery of the bore 50, it is apparent that the force P is ineffective to distort the bearing member 28 radially outward at those points. Thus,

the points 54,56 define points of minimum distortion while the points 90 on either side define points of potential maximum distortion. The result is that, under operating pressure, an elliptical configuration of the bearing member 28 is obtained. Since the bearing member 28 has distorted into an elliptical configuration, it will be appreciated that the ring gear 20, which has less structural rigidity, also will be forced to assume an elliptical configuration conforming to that of the bearing.

However, it is important to note that the major diameter of the ellipse is approximately at right angles to the neutral axis of the device. This relationship is of the utmost importance since, as pointed out above, it is the elliptical configuration with the major diameter coincident with the neutral axis which adversely affects operation of the device. By having the major diameter of the ellipse at right angles to the neutral axis of the device there is no adverse affect and, in fact, some definite advantages are obtained.

More specifically, since the major diameter of the ellipse lies at right angles to the neutral axis of the device, it follows that the minor diameter lies along the neutral axis. While the portions of the ring gear lying along the major diameter tend to move radially away from each other, the portions of the ring gear lying along the minor diameter of the ellipse tend to move radially toward each other. The result is closer meshing engagement of the gear teeth at the closed mesh position and an improved sealing relationship between the gear teeth at the open mesh position. Thus, instead of the distortion of the ring gear destroying the sealing engagement at the open mesh condition, it is apparent that, by controlling the distortion, an improved sealing relationship actually can be obtained.

Another important advantage of the controlled distortion of the ring gear is the improved performance which may be achieved. Thus, as the portions of the ring gear along the major diameter move outwardly it will be apparent that the chambers 32L will increase in volume. In other words, with the periphery of the pinion gear being fixed and distortion of the ring gear causing the ring gear to move radially outwardly away from the periphery of the pinion gear, the resulting fluid chamber defined between the teeth of the two gears is somewhat larger in volume than when the ring gear is not distorted under pressure. With this increased volume in chambers 32L, it is apparent that a greater volume of fluid may be received therein, thereby increasing the displacement per revolution of the pump. Thus, it is possible for the pump to displace a greater volume of fluid at high pressures than it i does at low pressures. This volumetric increase, at pressure, coupled with enhanced sealing at the open mesh condition, improves significantly the volumetric efficiency of a unit. For example, experimental results have indicated that, with this approach, a volumetric efficiency of 102 percent may be achieved at 2,000p.s.i..

The described mechanical arrangement is dependent on the generated pressures of the device for successful operation. It should be apparent that other mechanical approaches which might be independent of pressure could be employed.

MECHANICAL-HYDRAULIC DISTORTION CONTROL Referring now to the embodiment of FIG. 6, a modified form of the hydraulic device is illustrated. The modification of FIG. 6 is, in all respects, the same as the embodiment of FIG. I with the addition of controlled communication of high and low pressures at spaced places in the journal bearing clearance between the ring gear 20 and the bearing member 28. Thus, a pair of high-pressure grooves 60 are formed in the bearing surface 30 of the bearing member 28 at points spaced approximately 180 apart. Each of the grooves 60 communicates with the high-pressure manifold 42 through feed grooves 61 in wall 16. Spaced approximately 90 from the two high-pressure grooves, is a single low pressure or suction groove 62 which also is on the inner periphery of the bearing member 28 but which communicates through feed groove 63 with the lowpressure manifold 40.

It will be appreciated that the high pressure present in the grooves 60 result in hydraulic forces acting radially inward against the outer periphery of the ring gear 20, thereby assisting the bearing member 28 in collapsing the ring gear inwardly along the neutral axis. The low pressure or suction groove 62 defines a point of low pressure and thus encourages the mechanical distortion of the ring gear 20 radially outwardly in the area which is intended to correspond to the major diameter of the desired elliptical configuration. Thus, in the embodiment of FIG. 6, the combined hydraulic and mechanical forces achieve the desired result.

The grooves 60 could be positioned substantially in line with the neutral axis 34 but, as shown in FIG. 6, these grooves are spaced a few degrees before and after the neutral axis 34. The circumferential spacing of the grooves 60 from the neutral axis is desirable to counteract the directional rotation of the ring gear which carries the pressure in the direction of rotation. For example, at 2,000 rpm, an angular offset of 18 is approximately the correct location of the grooves with a journal clearance of 0.002 inches.

HYDRAULIC DISTORTION CONTROL In many applications, it may be satisfactory to dispense with I the mechanical forces generated by the special bearing member 28 and rely only on hydraulic forces to control distortion of the ring gear 20. In other words, the device may be the same as that shown in FIG. 6 except that the special bearing member is replaced by a conventional bearing.

To fully understand the nature of the distortion control with such an arrangement, it is helpful to consider a diagram of the hydraulic forces present in device of FIG. 6. Thus, referring to FIG. 8, there is schematically illustrated the ring gear 20 within a bearing member with the clearance between the ring gear and the bearing member being greatly exaggerated for purposes of illustration. As illustrated, the resultant force P acts against the inner periphery of the ring gear 20 to move the ring gear to the left as viewed in FIG. 8. A corresponding and opposite force Pj will be developed by the journal bearing to balance the gear 20 within the journal bearing. There also will be pressure gradients across the lands 45 and which will include points of half pressure (l/2P) acting radially outwardly substantially along the neutral axis of the device in the manner described above in connection with FIG. 4. In addition to these forces, the hydraulic pressure fed to the grooves 60 is the full hydraulic pressure of the device which will act radially inwardly along the neutral axis of the device. Since the suction groove 62 also communicates with the bearing clearance, it will be appreciated that the area within the clearance immediately adjacent the groove 62 will be substantially at zero pressure with a pressure gradient between the two grooves.

The resolution of the forces in FIG. 8 results in the configuration illustrated in FIG. 9. Thus, the forces P and P acting at right angles to the neutral axis of the device are equal and opposite. The resolution of the forces along the neutral axis of the device indicates that there will be radially inwardly directed forces tending to collapse the ring gear inwardly along the neutral axis. This then will force the ring gear to distort and go elliptical or egg shaped in configuration with the major diameter of the ellipse lying at right angles to the neutral axis, as shown in FIG. 9.

This then is the primary concept of the selective introduction of hydraulic pressures in the bearing clearance to control the ring gear distortion. However, a further aspect of this concept should be noted. Referring again to FIG. 9, and assuming the distortion of the ring gear 20 to oval shape, it is to be noted that the ring gear tends to cover the suction groove 62 and close it off. With the groove 62 restricted and the grooves 60 open, a condition of high pressure will be established in the bearing clearance almost completely around the outer periphery of the ring gear and, if it be assumed that the groove 62 is completely closed off, the pressure would reach a stabilized condition which would, in effect, cause the ring gear to return to its substantially cylindrical configuration. Of course, as soon as the ring gear 20 begins to return to a round configuration, the groove 62 becomes uncovered and the unbalanced forces which caused the gear to distort in the first instance are reestablished thereby returning the ring to an elliptical configuration.

From the foregoing, it will be appreciated that, by controlling the size of the suction groove 62, the amount of distortion of the gear can also be controlled. Thus, with a narrow suction groove, the ring gear will tend almost to completely close off the suction groove and thus establish, temporarily, a uniformity of pressure within the bearing clearance which will tend to return the ring gear to a cylindrical form. However, with a slightly wider suction groove, the ring gear can not completely close it off with the result that the imbalance of forces which caused the elliptical configuration which the ring gear assumes during operation. In this manner, the desired sealing action between the teeth at the open mesh condition and the proper driving relationship between the teeth at the closed mesh condition may be assured.

Having thus described my invention, I claim:

1. In a high-pressure hydraulic device having a housing, a cavity in the housing, inlet and outlet ports in the housing communicating with the cavity, a rotatable shaft extending into the cavity an element secured to said shaft for rotation therewith, a cylindrical member cooperating with said element to define increasing and decreasing volume fluid chambers therebetween, and a bearing supporting said cylindrical member in said housing and normally defining a bearing clearance therewith of from 0.0005 to 0.001 inches per inch of diameter, said hydraulic device having a neutral axis defined by a line drawn through the points of maximum and minimum volume within said fluid chambers, the improvement comprising:

means for deforming said cylindrical member to a generally elliptical configuration having its major diameter generally at right angles to said neutral axis.

2. The improvement of claim 1 wherein said deforming means comprise means responsive to the hydraulic pressures generated within the hydraulic device.

3. The improvement of claim 1 wherein said deforming means comprises a combination of hydraulic and mechanical forces acting on said cylindrical member.

4. The improvement of claim 1 wherein said deforming means comprises said bearing supporting said cylindrical member having a surface with a generally elliptical configuration having a major axis generally at right angles to said neutral axis.

5. The improvement of claim 4 and further including passage means communicating predetermined places on the outer periphery of said cylindrical member with high and low pressures.

6. The improvement of claim 5 wherein said low-pressure place is at generally right angles to and on the low pressure side of the neutral axis of the device and said high-pressure places are on said neutral axis.

7. The improvement of claim 4 wherein said bearing is a bearing member received in said cavity with the periphery of said bearing member so engaging the periphery of said cavity only at certain predetermined areas generally adjacent to the neutral axis that its bearing surface has a generally elliptical configuration which has its major diameter generally at right angles to said neutral axis.

8. The improvement of claim 7 and further including passage means communicating the high hydraulic pressures of said device to predetermined points on the outer periphery of said cylindrical member.

9. The improvement of claim 8 wherein said pressures are applied at points generally along the neutral axis of said device.

10. The improvement of claim 1 wherein said deforming means comprises high hydraulic pressure communicated to selected points generally on the neutral axis and low hydraulic pressure communicate to another selected point generally at right angles to the neutral axis.

11. The improvement of claim 10 wherein said high hydraulic pressures are applied to the outer periphery of said cylindrical member at points generally along the neutral axis of said device.

12. The improvement of claim 11 wherein a portion of the periphery of said cylindrical member spaced approximately from the neutral axis is in communication with a source of low hydraulic pressure.

13. The improvement of claim 1 wherein said cavity is generally cylindrical and said bearing member before insertion into said cavity has a cylindrical inner bearing surface and an outer diameter on said neutral axis greater than the diameter of said cavity and a lesser diameter elsewhere whereby when said bearing member is inserted into said cavity its bearing surface is forced to an elliptical configuration having its major axis generally perpendicular to said neutral axis.

14. The improvement of claim 13 wherein said diameter is greater at points spaced approximately 200 measured over the low-pressure side of said device.

15. The improvement of claim 13 wherein said bearing member has points on its bearing surface generally adjacent said neutral axis communicated with the high hydraulic pressure and a point generally intermediate said high-pressure points and on the low-pressure side of the device communicated with low pressure. 

1. In a high-pressure hydraulic device having a housing, a cavity in the housing, inlet and outlet ports in the housing communicating with the cavity, a rotatable shaft extending into the cavity an element secured to said shaft for rotation therewith, a cylindrical member cooperating with said element to define increasing and decreasing volume fluid chambers therebetween, and a bearing supporting said cylindrical member in said housing and normally defining a bearing clearance therewith of from 0.0005 to 0.001 inches per inch of diameter, said hydraulic device having a neutral axis defined by a line drawn through the points of maxiMum and minimum volume within said fluid chambers, the improvement comprising: means for deforming said cylindrical member to a generally elliptical configuration having its major diameter generally at right angles to said neutral axis.
 2. The improvement of claim 1 wherein said deforming means comprise means responsive to the hydraulic pressures generated within the hydraulic device.
 3. The improvement of claim 1 wherein said deforming means comprises a combination of hydraulic and mechanical forces acting on said cylindrical member.
 4. The improvement of claim 1 wherein said deforming means comprises said bearing supporting said cylindrical member having a surface with a generally elliptical configuration having a major axis generally at right angles to said neutral axis.
 5. The improvement of claim 4 and further including passage means communicating predetermined places on the outer periphery of said cylindrical member with high and low pressures.
 6. The improvement of claim 5 wherein said low-pressure place is at generally right angles to and on the low pressure side of the neutral axis of the device and said high-pressure places are on said neutral axis.
 7. The improvement of claim 4 wherein said bearing is a bearing member received in said cavity with the periphery of said bearing member so engaging the periphery of said cavity only at certain predetermined areas generally adjacent to the neutral axis that its bearing surface has a generally elliptical configuration which has its major diameter generally at right angles to said neutral axis.
 8. The improvement of claim 7 and further including passage means communicating the high hydraulic pressures of said device to predetermined points on the outer periphery of said cylindrical member.
 9. The improvement of claim 8 wherein said pressures are applied at points generally along the neutral axis of said device.
 10. The improvement of claim 1 wherein said deforming means comprises high hydraulic pressure communicated to selected points generally on the neutral axis and low hydraulic pressure communicated to another selected point generally at right angles to the neutral axis.
 11. The improvement of claim 10 wherein said high hydraulic pressures are applied to the outer periphery of said cylindrical member at points generally along the neutral axis of said device.
 12. The improvement of claim 11 wherein a portion of the periphery of said cylindrical member spaced approximately 90* from the neutral axis is in communication with a source of low hydraulic pressure.
 13. The improvement of claim 1 wherein said cavity is generally cylindrical and said bearing member before insertion into said cavity has a cylindrical inner bearing surface and an outer diameter on said neutral axis greater than the diameter of said cavity and a lesser diameter elsewhere whereby when said bearing member is inserted into said cavity its bearing surface is forced to an elliptical configuration having its major axis generally perpendicular to said neutral axis.
 14. The improvement of claim 13 wherein said diameter is greater at points spaced approximately 200* measured over the low-pressure side of said device.
 15. The improvement of claim 13 wherein said bearing member has points on its bearing surface generally adjacent said neutral axis communicated with the high hydraulic pressure and a point generally intermediate said high-pressure points and on the low-pressure side of the device communicated with low pressure. 